Continuously variable transmission

ABSTRACT

A continuously variable transmission employing a flat belt extending between driving and driven pulley assemblies, each having a circumferential array of radially adjustable belt engaging elements, is disclosed in conjunction with a control system. The radial positions of the belt engaging elements of each pulley assembly depend upon the angular relationship between inner and outer guideway disk structures carrying logarithmic spiral guideways oriented in the opposite sense. Each pulley assembly includes a power consuming or producing element, such as an oil pump/motor, which is coupled through harmonic drive differential gearing to the guideway disks. When the load on the hydraulic unit operated as a pump is changed by changing the output pressure against which it pumps, the angular relationship between the inner and outer guideway disk structures changes to correspondingly change the positions of the guideway intersections. Since the belt is of fixed length, the other pulley assembly must change its effective diameter in the opposite direction, and this response is facilitated by changing the output pressure of its oil pump in the opposite direction. An alternate configuration and control is disclosed for use when large pulley actuator torques are present due to the centrifugal forces acting upon heavier, stronger belt engaging elements selected for heavier duty applications.

CROSS REFERENCE TO RELATED APPLICATION

This application is a continuation-in-part of U.S. patent applicationSer. No. 051,922, filed May 19, 1987, by Emerson L. Kumm and alsoentitled CONTINUOUSLY VARIABLE TRANSMISSION, now U.S. Pat. No.4,768,996.

FIELD OF THE INVENTION

This invention relates to the continuously variable transmission (CVT)art and, more particularly, to a hydraulic/mechanical control system forestablishing the speed ratio in a flat belt continuously variabletransmission.

BACKGROUND OF THE INVENTION

Continuously variable transmissions of the class broadly characterizableas that in which a belt couples a pair of pulleys, each of which canassume a more or less continuous range of effective diameters, generallyfall into two categories; viz: (a) those employing V-belts or variationsthereof (such as link belts or chains) for transmitting power from onepulley to the other, and (b) those systems employing flat, flexiblebelts between the variable diameter pulleys.

Those skilled in the art have come to appreciate the CVT's employingflat, flexible belts enjoy significant fundamental advantages over thosesystems employing V-belts. In the case of the latter, the belts arecomposed of various compositions and have a trapezoidal cross section,the belt transmitting rotary motion at one speed from a source of power(such as an engine or motor) to an output shaft at another speed, thespeed ratio being varied in a continuous fashion from a minimum to amaximum as dependent on the geometry of the belt and the pulley system.The V-belt is compressed between smooth, conical sheave sections in eachof the two pulleys by external axial forces acting on the sections toapply tension or compression to the belt and friction between the sidesof the V-belt in the sheave sections to prevent slippage. In operation,a force unbalance caused by changes in the axial loading of the sheavesections causes the V-belt to change its radial positions in the twopulleys until a force balance is achieved or a limit range stop isreached.

For a large transmitted torque, the required axial forces exerted on thesheaves result in large compressive forces on the V-belt which requiresthat the belt have a substantial thickness to prevent its axial collapseor failure. This increase in thickness increases the belt's centrifugalforce and causes higher belt tension load. In addition, as the beltthickness increases, the pulley size must be increased due to higherstress loads at a given design minimum pulley radius. Further, thetypical V-belt must continuously "pull out" from the compressive sheaveload on leaving each pulley which results in significant friction lossesand belt fatigue which adversely affects the overall efficiency andoperating life. Consequently, although variable speed pulley drives havesuccessfully used V-belts in a wide range of applications, they havebeen severely limited in their power capabilities for more competitivesmaller size equipment.

As a result of these inherent drawbacks to the use of V-belts incontinuously variable transmissions, a second category has developedwhich may broadly be designated as flat belt drive continuously variabletransmissions. As the name suggests, flat belts are employed betweendriven and driving pulleys which are dynamically individually variablein diameter to obtain the sought-after ratio changes. No axial movementbetween the two pulley rims is necessary. On the other hand, it isnecessary to somehow effect the diametric variations of the individualpulleys, and in one particularly effective system, this function isachieved by causing a circular array of drive elements in each pulley totranslate radially inwardly or outwardly in concert as may beappropriate to obtain a given effective diameter of the pulley. Variablespeed flat belt transmissions of this particular type, and theirassociated control systems, are disclosed in U.S. Pat. Nos. 4,024,772;4,295,836 and 4,591,351 as well as U.S. patent application Ser. No.871,254 filed June 6, 1986, and now U.S. Pat. No. 4,714,452, all toEmerson L. Kumm. In all but the first patent enumerated above, thesubject variable diameter pulley components have included a pair ofpulley sheaves between which extend a series of belt engaging elementsthat are simultaneously moved both radially and circumferentially. Inone exemplary construction, there is a series of twenty-four beltengaging elements such that an angle of fifteen degrees extends betweenruns of the belt coming off tangentially from one belt engaging elementcompared to that of an immediately adjacent belt engaging element.

Each pulley sheave includes two pairs of two disks (designated,respectively, the inner guideway disk and the outer guideway disk ineach pair) which are parallel to each other with the inner and outerguideway disks of each pair being disposed immediately adjacent oneanother. Each of the guideway disks of an adjacent pair has a series ofspiral grooves or guideways with the guideways of the pair oriented inthe opposite sense such that the ends of the belt engaging elements arecaptured at the intersections of the spiral guideways. Thus, the radialadjustment to the belt engaging elements may be achieved by bringingabout relative rotation between the inner and outer guideway disks tochange their angular relationship, this operation being, of course,carried out simultaneously and in coordination at both pairs of guidewaydisks of a pulley. Thus, it will be appreciated that the overalltransmission ratio is dependent upon the respective angularrelationships between the facing pairs of guideway disks on each of thetwo pulleys.

From the foregoing, it will be understood that the control system whichestablishes the instantaneous angular relationship between the facingdisks of each pair on each of the pulleys is a highly important systemwithin the entire continuously variable transmission. A series ofrelated disadvantages has been characteristic of the prior art controlsystems for establishing these angular relationships. More particularly,these disadvantages include the fact that the mechanical components ofthe prior art hydraulic/mechanical control systems have been physicallylarge and heavy, contributing the majority of the overall size of thesecontinuously variable transmissions and accounting for a large portionof their weight. Also, the prior art control systems have requiredsubstantially larger hydraulic supply pressures to achieve the desiredcontrol response rates. As a result, the hydraulic control systems havebeen, of necessity, correspondingly complex, further contributing to thesize, weight and cost of the prior art continuously variabletransmissions.

It is to provide a control system which overcomes these several relatedobjections to the prior art control systems for flat belt continuouslyvariable transmissions that my invention is directed.

OBJECTS OF THE INVENTION

It is therefore a broad object of my invention to provide an improvedflat belt continuously variable transmission.

It is another object of my invention to provide, in such a continuouslyvariable transmission, an improved control system for establishing theeffective transmission ratio between an input shaft and an output shaftcoupled by a flat belt.

It is a still further object of my invention to provide such a controlsystem which is relatively small, lightweight, simple and inexpensive.

It is yet another object of my invention to provide a continuouslyvariable transmission in which the transmission speed ratio may be morerapidly changed than in the past.

In another aspect, it is a more specific object of my invention toprovide such a control system which is self-energized from the energytransmitted to the pulley.

Still more specifically, it is an object of my invention to provide sucha control system in which the effect of centrifugal forces exerted byrelatively heavy belt-engaging elements are used to advantage andcompensated for as necessary.

SUMMARY OF THE INVENTION

Briefly, these and other objects of my invention are achieved byproviding, in a flat belt continuously variable ratio transmission, acontrol system which obtains the energy necessary to effect pulleydiameter changes from the energy source driving the pulley and beltassembly rather than from an external source. The inside and outsideguideway disks on both sides of each pulley are connected together usingdifferential gearing which is also coupled to a power consuming element.While it is possible to use many different arrangements of differentialgearing to obtain the desired operating torques between the inner andouter guideway disks, the presently preferred embodiment of theinvention employs a harmonic gear drive to provide the differentialgeared relationship between the inner and outer guideway disks and apower consuming element. The harmonic drive consists of an ellipticallyshaped wave generator and a flexible splined gear sleeve (theflexspline) geared to two circular, internally splined rings designatedthe dynamic spline and the circular spline, respectively. The circularspline typically has two more splines or teeth than the dynamic splineand the flexspline (which have the same number) resulting in a speedreduction between the wave generator and the dynamic spline equal toone-half of the number of splines on the dynamic spline. Hence, with theharmonic drive, it is possible to obtain speed reductions or torqueincreases from the wave generator to the circular spline or dynamicspline exceeding 100 to 1 in a single stage of gearing.

In the preferred embodiment of the continuously variable transmissiondisclosed in the above referenced parent patent application Ser. No.051,922, for the driven pulley, the circular spline is connected to theinner guideway disks and the dynamic spline is connected to the outerguideway disks. Conversely, for the driving pulley (which is connectedto the external power source such as a motor or engine), the dynamicspline is connected to the inner guideway disks and the circular splineis connected to the outer guideway disks. The orientation of theguideway disks as related to the direction of rotation of the pulleysmust be as specified in the previously referenced U.S. patentapplication Ser. No. 871,254, now U.S. Pat. No. 4,714,452. With thisarrangement, a drag torque provided by an output power consuming elementconnected to the wave generator of each pulley assembly gives asubstantially proportional pulley actuator torque which, working throughthe harmonic gear drive, tends to move the belt engaging elements of thepulley radially outwardly. Hence, a fixed length belt can be tensionedaround two rotating pulleys for transmitting torque using a drag torqueon the output power consuming element in each pulley to achieve aself-energized actuator drive. Since very high gear ratios may be used,correspondingly small torques (and hence small powers) are transferredto the output power consuming element to thereby permit the generationof large actuator torques to be applied for maintaining a constant speedratio in the CVT or, when desired, to change the CVT speed ratio by anappropriate transient application to dynamically adjust the angularrelationships between the inner and outer guideway disks for eachpulley.

While numerous arrangements can be used for the control torqueabsorption on the output power consuming element (including anelectrical power generator, a friction clutch, an air compressor,variable viscous fluid coupling, etc.), the employment of a positivedisplacement oil gear pump for this operation permits a wide range ofdrag torques to be effected simply by opening and closing an outputvalve to variably load the oil pump. In a refinement of the controlsystem, the oil pumps may selectively transiently operate as motorsduring rapid ratio changes. An oil supply subsystem is correspondinglytransiently actuated to facilitate this temporary mode of operation.

In the disclosure of the parent application, the use of relativelylightweight belt-engaging elements was assumed with such materials asaluminum, ceramic and plastic contemplated. However, it is desirable toconsider the use of more dense belt engaging elements (i.e., those madeof steel, powdered steel alloys, etc.), which have fatigue strengthsmuch higher than aluminum, ceramic or plastic, as the design pulleypower is increases. When such relatively dense belt drive elements areemployed, the centrifugal forces caused by pulley rotation and tendingto counter rotate the two sets of discs in each pulley in the directionthat moves the belt engaging elements radially outwardly becomedominant, and these centrifugal forces can be used to advantage. Thus,at high rotational pulley speeds, the torques generated by thecentrifugal forces acting on the heavier belt engaging elements may bebalanced by the torque of a hydraulic unit through the differentialgearing of a harmonic drive if the specific connections between thedynamic spline or circular spline and the inner or outer guideway discsare reversed from those employed with the lighter weight belt engagingelements. In addition, for rotational speeds below those at which thecentrifugal forces are dominant, the hydraulic units may be operated asmotors, under the influence of a revised hydraulic control subsystem,to, in effect, return the CVT to the same mode of operation employedwith the lighter weight belt engaging elements.

DESCRIPTION OF THE DRAWING

The subject matter of the invention is particularly pointed out anddistinctly claimed in the concluding portion of the specification. Theinvention, however, both as to organization and method of operation, maybest be understood by reference to the following description taken inconjunction with the subjoined claims and the accompanying drawing ofwhich:

FIG. 1 illustrates an edge on view of driving and driven pulleys coupledby a flat belt and representative of the class of continuously variabletransmissions in which the present invention finds application;

FIG. 2 is a cross sectional view, taken along the lines 2--2 of FIG. 1,of the pulley system illustrated in FIG. 1;

FIG. 3 is a fragmentary perspective view, partially broken away, of apulley particularly illustrating the relationships between inner andouter guideway disk components and belt engaging element components;

FIGS. 4a, 4b and 4c are illustrations showing the principle of operationof a harmonic drive, certain components being shown in an exaggeratedelliptical shape in order to more clearly demonstrate the principle;

FIG. 5 is a simplified cross sectional view of a flat belt continuouslyvariable transmission illustrating the fundamental aspects of themechanical components of the subject control system for establishing theangular relationship between the inner and outer guideway disks of eachpulley;

FIG. 6 illustrates a hydraulic control subsystem for use in conjunctionwith the mechanical control subsystem illustrated in FIG. 5 to establishthe angular relationship between inner and outer guideway disksaccording to the load and other demands;

FIG. 7. is a view similar to FIG. 5 illustrating a variant embodiment ofthe continuously variable transmission in which relatively heavy beltengaging elements are employed; and

FIG. 8 is a view similar to FIG. 6 and illustrating a hydraulic controlsystem suitable in conjunction with the mechanical control subsystemshown in FIG. 7.

DETAILED DESCRIPTION OF THE INVENTION

Referring now to FIGS. 1, 2 and 3, fundamental aspects of the flat belttype of a continuously variable transmission, with which the subjectcontrol system is employed, are illustrated as embodied in a variablediameter pulley drive assembly 10 comprising variable diameter pulleys11 and 12 connected by a flat drive belt 13. The pulley 11 will beconsidered as the driving pulley and the pulley 12 as the driven pulleyin this discussion, but it will be understood that the roles of thesepulleys may be reversed without altering the concepts involved.

The pulley 11 is appropriately mounted on a shaft 14, and the pulley 12is similarly appropriately mounted on a shaft 15 as is well understoodin the art. The pulleys 11 and 12 are similar to each other, and onlyone of them, namely 11, will be specifically described in thisdiscussion. The belt 13 as shown in FIG. 3 corresponds to the positionof the belt 13 of FIG. 2 in the dashed line position.

The pulley 11 includes a pair of pulley sheaves 16 and 17 between whichextend a series of belt engaging elements 18, the latter being engagedby the belt 13 for driving, or driven, conditions as will be understood.In one construction of the invention, there is a series of twenty-fourbelt engaging elements 18 equally circumferentially distributed wherebythere is established an angle of fifteen degrees between runs of thebelt 13 coming off tangentially from one belt engaging element 18 ascompared to that of an immediately adjacent belt engaging element 18.Each belt engaging element 18 includes a central shank 28, which engagesthe belt 13, and bearing regions 29 at each end.

The pulley sheave 16 comprises a pair of pulley guideway disks 19 and 21which are parallel to and lie immediately adjacent each other injuxtaposition. Similarly the pulley sheave 17 comprises a pair of pulleyguideway disks 22 and 23 which are parallel to and lie immediatelyadjacent each other in juxtaposition. The longitudinal spacing betweenthe pulley sheaves 16 and 17 (i.e., the axial spacing between guidewaydisks 21 and 22) remains the same irrespective of the radial adjustmentof the belt 13 for different driving or driven speeds. This spacing issufficient to accommodate with clearance the width of belt 13 which isselected to carry the load that the system is designed for as is wellunderstood.

The range of radial adjustment or position of the belt 13 on the pulley11, as may be envisioned by the solid line and dashed line positions ofbelt 13 in FIG. 2, is achieved by altering the radial positions of thebelt engaging elements 18. For example, in FIG. 2 the belt engagingelements 18 are close to the center of the shaft 14 in the solid lineposition of the belt 13 on pulley 11; conversely, the belt engagingelements are radially farther out, namely adjacent the periphery, whenthe belt 13 is in its dashed line position which is also the positionshown in FIG. 3.

Variation in the radial positions of the belt engaging elements 18 isachieved by relative rotation to change the angular relationship betweenthe outer guideway disk 19 and the inner guideway disk 21 of pulleysheave 11 and by identical relative rotation of the pulley guidewaydisks 22 and 23 of pulley sheave 17. As a practical matter, to insuresynchronous operation, the inner guideway disks 21 and 22 are physicallylocked together, and the outer guideway disks 19 and 23 are also lockedtogether. Power for such operation, not shown in FIGS. 1, 2 or 3, hasbeen achieved in the prior art typically as disclosed in U.S. Pat. No.4,295,836 previously referenced.

The outer guideway disk 19 has a series of logarithmic spiral guideways24 therein which progress outwardly from adjacent the center at an angleof forty-five degrees with respect to the pulley radius. Simiarly theinner guideway disk 21 has a series of logarithmic spiral guideways 25radiating outwardly at an angle of forty-five degrees with respect tothe pulley radius, but in the opposite sense to the guideways 24 ofpulley disk 19. Since the guideways 24 and 25 radiate outwardly atangles of forty-five degrees with respect to the pulley radius, but inopposite senses, the intersections of these guideways exist at ninetydegrees at all radial positions. This results in a substantiallyconstant geometry at the intersections of the logarithmic spiralguideways 24 and 25 at all radial positions for receiving the bearingregion ends 29 of the belt engaging elements 18. Similarly, the innerguideway disk 22 has a series of logarithmic spiral guideways 26radiating outwardly identically to the guideways 25 of inner guidewaydisk 21, and the outer guideway disk 23 includes logarithmic spiralguideways 27 extending outwardly identically to the guideways 24 ofouter guideway disk 19. Hence, the guideways 26 and 27 intersect atninety degrees at all radial positions to give a constant intersectiongeometry identical to the logarithmic spiral guideways 24 and 25 forreceiving the other ends of the belt engaging elements 18.

While forty-five degree spirals have been shown and give ninety degreeintersections, it will be understood that logarithmic spirals of otherangularities may be used as desired. Also, minor variations from aparticular angularity may be tolerated so long as the belt engagingelement bearing ends supported at the guideway intersections will moveappropriately when the sheaves are rotated relative to each other tochange the angular relationship between the inner and outer guidewaydisks.

It will be clear that the belt 13, as it passes around the pulley 11 or12, engages the central portion of the belt engaging elements 18 andcauses one pulley to drive and the other pulley to be driven in theobvious fashion.

The foregoing description of the basic drive system, the pulleys 11 and12, the belt 13 and the belt engaging elements 18 is set forth ingreater detail in U.S. Pat. No. 4,295,836, dated Oct. 20, 1981,previously referred to and does not form a specific part of theinvention described in this application, but forms the environment inwhich the invention functions.

While it is possible to use many different differential gearingarrangements to obtain the desired operating torques between the innerand outer guideway disks of the pulleys (which, as will be seen below,is a necessary operation for practicing the present invention), thepresently preferred embodiment employs a so-called "harmonic" gear driveto provide the differential geared relationship between the inner andouter guideway disks and certain power absorbing elements. Referring toFIGS. 4a, 4b and 4c, and particularly to the somewhat enlarged FIG. 4a,the basic principles of a harmonic drive gear reduction apparatus arepresented. In this most elementary form, a harmonic drive employs threeconcentric components to produce high mechanical advantage and speedreduction. The use of nonrigid body mechanics allows a continuouselliptical deflection wave to be induced in a nonrigid external gear,thereby providing a continuous rolling mesh with a rigid internal gear.Thus, referring to FIG. 4a, an elliptical wave generator 30 deflects aflexspline 31 which carries outside teeth and therefore meshes with theinside teeth of a rigid circular spline 32. The elliptical shape of theflexspline and the amount of flexspline deflection is shown greatlyexaggerated in FIGS. 4a, 4b and 4c in order to demonstrate theprinciple. The actual deflection is very much smaller than shown and iswell within the material fatigue limits.

Since the teeth on the non-rigid flexspline 31 and the rigid circularspline 32 are in continuous engagement and since the flexspline 31typically has two teeth fewer than the circular spline 32, onerevolution of the wave generator 30 causes relative motion between theflexspline and the circular spline equal to two teeth. Thus, with thecircular spline 32 rotationally fixed, the flexspline 31 will rotate inthe opposite direction to the wave generator (the system input in theexample) at a reduction ratio equal to the number of teeth on theflexspline divided by two.

This relative motion may be visualized by examining the motion of asingle flexspline tooth 34 over one-half of an input revolution in thedirection shown by the arrow 35. Since the input to the wave generator30, in the example, causes clockwise rotation of the wave generator, theflexspline rotates counterclockwise. Thus, referring to FIG. 4b, it willbe seen that the tooth 34, after one-quarter revolution of the wavegenerator 30, has moved counterclockwise one-half of one flexsplinetooth position. It will also be noted that when the wave generator 30axis has rotated 90°, the tooth 34 is fully disengaged. Fullreengagement occurs in the adjacent circular spline tooth space when themajor axis of the wave generator 30 is rotated to 180° as shown in FIG.4c, and the tooth 34 has now advanced one full tooth position. Thismotion repeats as the major axis rotates another 180° back to zero,thereby producing the two tooth advancement per input revolution to thewave generator 30.

Conventional tabulations of harmonic drive gear reduction ratios assumethe flexspline is the output member with the circular splinerotationally fixed. However, any of the drive elements may function asthe input, output or fixed member depending upon whether the gearing isused for speed reduction, speed increasing or differential operation.

The harmonic drive principle can be extended by the addition of a fourthelement designated the dynamic spline. The dynamic spline is an internalgear that rotates at the same speed and in the same direction as theflexspline. Unlike the circular spline (to which it is parallel, alsoengaging the flexspline), the dynamic spline has the same number ofteeth as the flexspline. Flexspline shape rotation results in toothengagement/disengagement within the same tooth space of the dynamicspline such that the ratio between the two is one to one. The system,therefore, is a flexspline output with the same characteristics as thethree element harmonic drive model; i.e., gear reduction ratio tabulatedwith the direction of rotation opposite to the input. Ultra high dualratio capability can be obtained by using two circular splines in meshwith the flexspline with each developing a different single-stage ratio.Merely by way of example, the compounding of single-stage ratios of160:1 and 159:1 results in a total reduction ratio of 12,720:1. Harmonicdrives suitable fo use in the present invention may be obtained from theHarmonic Drive Division of Emhart Machinery Group in Wakefield, Mass.

The subject invention, in the presently preferred embodiment, employs afour element harmonic drive in which the fourth element is a dynamicspline. Referring now to FIG. 5, a slightly simplified representation ofa flat belt continuously variable transmission according to the presentinvention is shown. Preliminarily, it may be noted that, with the properconnection of the dynamic spline or the circular spline to the innerguideway disks or the outer guideway disks on a given rotating pulleyusing the guideway disk orientation as related to the direction ofpulley rotation and direction of power flow as given in the previouslyreferenced U.S. patent application Ser. No. 871,254 (now U.S. Pat. No.4,714,452), a drag torque on a power consuming element connected to thewave generator can be caused to give a largely proportional force whichtends to move the belt engaging elements radially outwardly. As aresult, a fixed length belt can be tensioned around two rotating pulleysfor transmitting torque using a drag torque on the energy consumingelement in each pulley to give a self-energized actuator drive. Sincevery high ratios (greater than 100:1) may be used, very small torques(and hence small powers) transmitted to the power absorbing elementachieves the generation of large actuator torques to be applied inpositioning the belt engaging elements of the continuously variabletransmission pulleys.

In FIG. 5, the pulley assembly 40 may be deemed the driving pulley(which receives torque via the input shaft 42 from an external powersource not shown such as an engine or motor) and the pulley assembly 41may be deemed the driven pulley which receives power via the flat belt43 which is applied to an output shaft 44. As may be appreciated fromthe manner in which FIG. 5 is cross hatched, the inner guideway disks 45of the pulley assembly 40 are physically connected together to effect aninner guideway disk structure. Simlarly, the outer guideway disks 46 ofthe pulley assembly 40 are also connected together to effect an outerguideway disk structure. The inner guideway disks 47 and the outerguideway disks 48 of the driven pulley assembly 41 are similarlyconnected. Briefly comparing FIG. 5 to FIGS. 2 and 3, the innerguideways disks 45 correspond to the inner guideway disks 21, 22; theouter guideway disks 46 correspond to the outer guideway disks 19, 23;and the directions of rotation are as indicated by the arrows in eachFig. to establish the correct relationship between the components, asdiscussed more fully in the previously referenced U.S. patentapplication Ser. No. 871,254 (now U.S. Pat. No. 4,714,452), whichcontribute to the correct operation of the subject invention.

A four element harmonic drive 50 differentially interconnects the outerguideway disks 46, the inner guideway disks 45 (which are fastened tothe shaft 42), and an output drive 54 of the pulley assembly 40 to apower consuming element such as an oil pump 55. Similarly, as to thepulley assembly 41, a four element harmonic drive generator 51differentially connects the outer guideway disks 48, the inner guidewaydisks 47 (which are fastened to the shaft 44), and an output drive 56which is coupled to a second power consuming element such as an oil pump57.

In the driving pulley assembly 40, the inner guideway disks 45 areconnected to the dynamic spline 60 of the harmonic drive 50, and theouter guideway disks 46 are connected to the circular spline 61. Theinner guideway disks 45 are connected to the shaft 42 by a collar 52over one element of the outer guideway disks 46. The output drive 54between the oil pump 55 and the harmonic drive 50 is coupled to the wavegenerator 58.

The driven pulley assembly 41 is similarly configured, but the positionsof the dynamic spline and the circular spline are reversed. Thus, thecircular spline 62 is connected to the inner guideway disks 47, and thedynamic spline 63 is connected to the outer guideway disks 48. The innerguideway disks 47 are connected to the shaft 44 by collar 53 over oneelement of the outer guideway disks 48. The output drive 56 to the oilpump 57 is coupled to the wave generator 59 of the harmonic drive 51.

Any constant radial position for the belt engaging elements 64 (drivingpulley assembly 40) or 65 (driven pulley assembly 41) results in allcomponents of the harmonic drive for that pulley assembly (i.e., thewave generator, the flexspline, the dynamic spline, and the circularspline) rotating at the same speed as the pulley shaft. Hence, the powerconsuming element (the oil pumps 55 or 57 in FIG. 5) rotates at a speedproportional to the shaft speed producing a hydraulic oil flow whosepressure output (against which it works) can be changed by a controlvalve. It can be shown that the drag torque of the positive displacementoil pump used as the power consuming element is substantiallyproportional to the generated oil pressure. Thus, the actuator torque inthe pulley can be maintained constant at any belt drive radius andpulley speed by holding the positive displacement pump dischargepressure constant. Consequently, the pulley speed ratio may be changedby increasing the actuator torque on one pulley versus the other.

For example, if a higher output speed is desired for a given inputspeed, increasing the discharge pressure on the oil pump 55 would causethe wave generator 58 to transiently rotate more slowly relative to theshaft 42 causing the circular spline 61 connected to the outer guidewaydisks 46 to rotate more slowly relative to the inner guideway disks 45connected to the shaft 42. The movement of the inner guideway disksrelative to the outer guideway disks causes the intersections of theguideways, and hence the belt engaging elements 64, to move radiallyoutwardly in the driving pulley assembly 40. With a fixed length belt,this can only happen if the belt drive radius in the driven pulleyassembly 41 simultaneously decreases. An increase in belt tension due tothe increase in actuator torque in driving pulley assembly 40 willresult in increasing the torque on the power consuming unit, oil pump57, in the driven pulley assembly 41 subsequently increasing therotational speed of the oil pump 57 as the belt drive radius of drivenpulley assembly 41 is decreasing.

Torque can be transmitted in either direction through the harmonicdrives 50, 51 although the drive effficiency is a few percentage pointslower when transmitting power from the circular spline or dynamic splineto the wave generator as compared to the reverse case. This is not asubstantial concern in the subject system since operation at anyconstant speed ratio gives very low power losses in the transmission ofpower to the output power consuming elements (oil pumps 55, 57) since,again, there is no change in the relative position of any elements inthe pulleys or harmonic drives during operation at a constanttransmission ratio. When a speed change occurs, the temporary increasein power loss is inversely dependent upon the time that it takes for thespeed ratio change to be completed. Typically, the angular movement ofthe outer guideway disks relative to the inner guideway disks of a givenpulley for maximum radius ratio change (speed ratio change) is about100° of angular shift. Hence, if this change occurs in one second(typical of a maximum speed change), this corresponds to a 16.7 RPMspeed of the inner guideway disks relative to the outer guideway disksfor that period. Given an examplary 100:1 harmonic drive, a temporarychange in the power consuming unit of 1667 RPM is brought about.Increasing the rotational rate of the drive to the oil pump 57 by 1667RPM for one second will cause the output pulley assembly 41 to increasein speed and require a decrease in the oil pump 55 in the driving pulley40 by 1667 RPM for the same one second period using pulleys of the samesize.

The discharge oil pressure from the oil pump 55 in the driving pulleyassembly 40 may be increased by restricting its output control valveflow area to effect such a decrease in the oil pump 55 speed and providethe higher actuator torque to move the belt engaging elements 64radially outwardly. The simultaneous increase in the speed of the oilpump 57 of the driven pulley 41 can simultaneously be aided by openingits output control valve flow area.

In an automotive application for the continuously variable transmission,the output shaft 44 of the driven pulley assembly 41 is geared directlyto the vehicle wheels so that the rotational rate of the driven pulleyassembly is directly proportional to the vehicle speed. The inertia of avehicle does not permit the absolute value of the driven pulleyassembly's 41 speed to increase very rapidly (i.e., in a second or two).However, the more critical operation for an automotive continuouslyvariable transmission installation consists of obtaining maximum vehicleacceleration at any drive speed. This corresponds to rapidlyaccelerating the engine connected to the driving pulley assembly 40 to ahigher speed to give more power to accelerate the vehicle. In such acase, the discharge oil pressure from the oil pump 57 of the drivenpulley assembly 41 would be increased by restricting its output controlvalve flow area to effect a decrease in the oil pump 57 speed and givehigher actuator torque to move the belt engaging elements 65 radiallyoutwardly. The simultaneous increase in the speed of the oil pump 55 ofthe driving pulley assembly 40 can also be aided by opening its outputcontrol valve flow area. The overall resulting output torque from thecontinuously variable transmission versus time depends on many factors,but chiefly the inertia of the engine and the other components and thecontrolled output pressures of the oil pumps 55, 57.

The simultaneous rate of increase in the speed of the oil pump 55 or 57can be aided by operating the accelerating oil pump as a motor for abrief time duration (e.g., a second to a few seconds for any one speedchange). This feature can be incorporated by transiently supplying oilto the oil pump functioning as a motor from an independent boost pump,under control of an appropriate solenoid valve, for the very briefduration required to accommodate the rapid engine input speedacceleration or deceleration. The boost pump volume flow must beadequate for the maximum flow requirement, but the boost pressure can berelatively low thus keeping the motor power requirement to about 5-10%of the minimum oil pump power. The control system must always maintainadequate belt tension to prevent slippage during speed ratio changes.

One or more circumferential springs (represented schematically by thesprings 68, 69 in FIG. 5) may also be incorporated in each actuatordrive oriented to give a torque between the inner guideway disks andouter guideway disks whose direction would cause the belt engagingelements to be moved radially inwardly in the driving pulley assembly 40and radially outwardly in the output or driven pulley assembly 41. Thisarrangement permits the continuously variable transmission to startoperation at a maximum output torque to input torque ratio with the beltunder some tension at all times to avoid initial slippage. Such springsalso are helpful in obtaining a maximum speed increase of the drivingpulley assembly 40 relative to the driven pulley assembly 41 as desiredduring the acceleration of the vehicle; i.e., if the oil pressure on theoil pump 55 is reduced sufficiently, the circumferential spring torquemust help drive the belt engaging elements 64 radially inwardlypermitting very rapid engine acceleration. However, the specific desiredchanges as applied to the automotive continuously variable transmissionapplication normally require that there is no net loss in output torqueduring operator-demanded vehicle acceleration that has a duration ofmore than about 0.1 second. As a result, the rate and magnitude ofoutput pressure changes in the oil pumps 55, 57 have the major effect.

Consider now the exemplary hydraulic control subsystem which determinesthe input and output pressures of the oil pumps 55, 57 for the reasonspreviously discussed to control the speed ratio between the drivingpulley assembly 40 and the driven pulley assembly 41 via the belt 43,all as shown in FIG. 6. Oil from a reservoir 70 is supplied, throughconduit 71 and check valves 72, 73, to the suction sides 74, 75,respectively, of oil pumps 55, 57. The pressure sides 76, 77,respectively, of the oil pumps 55, 57 are connected, by respectiveconduits 78, 79, back to the reservoir 70 after the oil has passedthrough a cooler 80. The conduit 78 includes an inline pressure sensor81 and, downstream of the pressure sensor 81, a flow rate control valve82. Similarly, the conduit 79 includes an inline pressure sensor 83 anda flow rate control valve 84 downstream from the pressure sensor 83.

The status of the flow rate control valves 82, 84 is determined byoutputs from a control module 86 which receives input information fromthe pressure sensors 81, 83 and from external sources such as enginespeed and throttle position. When the control module senses that thepressures existing in the cnduits 78 and 79, from the sensors 81, 83(and hence the transmission ratio between the pulley assemblies 40, 41),are incorrect for sensed engine speed and throttle position conditions(as, for example, during rapid acceleration), the control moduleresponds by increasing the flow area through the flow rate control valve82 and decreasing the flow area through the flow rate control valve 84which momentarily slows the oil pump 57 and speeds the oil pump 55 untilthe effective radius of the driving pulley assembly 40 is decreased andthat of the driven pulley assembly 41 is increased to reach a new andcorrect system balance point. Also, when decelerating and using theengine as a partial vehicle brake, the control module 86 causes the flowrate through the flow rate control valve 84 to decrease and that throughthe flow rate control valve 82 to increase, momentarily slowing the oilpump 57 and accelerating the oil pump 55 until a system balance is againachieved.

As previously discussed, under certain conditions, the response speed ofthe oil pump 55 or 57 whose speed is being increased can be facilitatedby permitting it to operate briefly as a motor. This feature isaccomplished with the addition of a low pressure pump 88 driven by amotor 89. The pump 88 is supplied from the reservoir 70 and isconnected, via conduits 90, 91, 92 to the suction sides 74, 75 of theoil pumps 55, 57, respectively. A solenoid operated valve 93 in line inthe conduit 91, when opened under the influence of the control module86, supplies an extra volume of oil to the suction side 74 of the pump55 to permit its transient operation as a motor. Similarly, solenoidoperated valve 94, in line in the conduit 92, when opened under theinfluence of the control module 86, permits the pump 88 to supply anextra volume of oil to the suction side 75 of the pump 57 to permit itstransient operation as a motor. The pump 88/motor 89 and the solenoidoperated valves 93, 94 need only be operated when rapid transmissionratio changes are undertaken and only facilitate that rapid change.Thus, it will be appreciated that the portion of the hydraulic circuitincluding the pump 88, the conduits 90, 91, 92 and the solenoid operatedvalves 93, 94 is optional and may not be required for all systems.

The preceding description of the operation and specific connectionsbetween the inner guideway discs and dynamic spline or circular splineand also between the outer guideway discs and circular spline or dynamicspline for the input and output power pulleys assume the use ofrelatively light weight belt engaging elements. Such belt engagingelements may be made typically of aluminum, ceramic or plastic for lowweight. The belt engaging elements are acted upon by centrifugal forcecaused by pulley rotation and, due to their positions in the guidewaydiscs, give torques operating to counter rotate the two sets of discs ineach pulley in the direction that moves the belt engaging elementsradially outwardly. The effect of the centrifugal force acting on therelatively light weight belt engaging elements of each pulley is usuallyinsufficient to require provision for compensation in the controlsystem.

However, it is desirable to consider the use of more dense belt engagingelements (i.e., those made of steel, powdered steel alloys, etc.), whichhave fatigue strengths much higher than aluminum, ceramic or plastic, asthe design pulley power is increased. At high rotational pulley speeds,the torque generated by the centrifugal force of the heavier beltengaging elements may be balanced by the torque of a hydraulic pumpthrough the differential gearing of a harmonic drive provided only thatthe specific connections between the dynamic spline or circular splineand the inner or outer guideway discs are reversed from those shown inthe preceding description in which the use of relatively light weightbelt engaging elements was assumed.

Thus, referring now to FIG. 7 which illustrates a flat belt continuouslyvariable transmission employing belt engaging elements 164, 165 ofsignificant weight, the driving pulley assembly 140 has its outerguideway discs 146 connected to the dynamic spline 160, and the innerguideway discs 145 are connected to the circular spline 161. Similarly,the driven pulley assembly 141 has its outer guideway discs 148connected to the circular spline 162, and the inner guideway discs 147are connected to the dynamic spline 163. It will again be noted thatthese relationships are reversed from those shown and discussed withreference to the embodiment illustrated in FIG. 5.

As the operating speeds of the pulley assemblies 140, 141 decrease, thecentrifugal forces acting on the belt engaging elements 164, 165decrease with the square of the speed change. Thus, at lower pulleyassembly speeds, there is less torque on the guideway discs from thissource and correspondingly less tendency to cause the belt engagingelements to move radialy outwardly. Therefore, at lower speeds, it isnecessary to operate the hydraulic units 155, 157 connected to the wavegenerators 158, 159 of the harmonic drives as motors rather than aspumps in order to provide the necessary torque tending to move the beltengaging elements outwardly to obtain the required belt tension fortransferring power between the pulleys without slippage. In a givencontinuously variable transmission employing the heavier belt engagingelements, the pulley assembly speed below which it is necessary tooperate each of the hydraulic units 155, 157 as a motor rather than as apump is dependent chiefly on the weight of the belt engaging elements164, 165, the pulley torque and the overall effective friction factorbetween the belt 143 and the belt engaging elements. However, since thehydraulic flow rate to a fixed displacement hydraulic motor variesdirectly with its speed, the required auxiliary supply pump flow andpower required are minimized by operating only in the lower speed regimeof the pulleys.

The hydraulic control subsystem in this arrangement for using heavier,stronger belt engaging elements therefore requires an auxiliary pump tooperate continuously at the lower pulley assembly speeds to supply therequired hydraulic flows to the hydraulic units connected to the wavegenerators in the harmonic drives on the pulleys. (It may be noted thathydraulic flow requirements for a constant pulley torque will decreaseas the speed decreases.)

An exemplary hydraulic control subsystem for performing this function isillustrated in FIG. 8. While this embodiment of the hydraulic controlsubsystem bears a substantial resemblance to that illustrated in FIG. 6,many of the signals and subsystem responses are reversed as aconsequence of the differences in the mechanical configurationsillustrated in FIGS. 5 and 7 as discussed immediately above. However,the primary functions of the hydraulic control subsystem illustrated inFIG. 8 remains the determination of the input and output pressures ofthe hydraulic units 155, 157 to control the speed ratio between thedriving pulley assembly 140 and the driven pulley assembly 141 via thebelt 143.

Thus, oil from a reservoir 170 is supplied, through conduit 171 andcheck valves 172, 173, to the suction sides 174, 175, respectively, ofhydraulic units 155, 157. The pressure sides 176, 177, respectively, ofthe hydraulic units 155, 157 are connected, by respective conduits 178,179, back to the reservoir 170 after the oil has passed through a cooler180. The conduit 178 includes an inline pressure sensor 181 and,downstream of the pressure sensor 181, a flow rate control valve 182.Similarly, the conduit 179 includes an inline pressure sensor 183 and aflow rate control valve 184 downstream from the pressure sensor 183. Alow pressure pump 188, driven by a motor 189, is supplied from thereservoir 170 and is connected, via conduits 190, 191, 192 to thesuction sides 174, 175 of the hydraulic units 155, 157, respectively.

A solenoid operated valve 193 in line in the conduit 191, when openedunder the influence of the control module 186, supplies an extra volumeof oil to the suction side 174 of the hydraulic unit 155 to permit itsoperation as a motor. Similarly, solenoid operated valve 194, in line inthe conduit 192, when opened under the influence of the control module186, permits the pump 188 to supply an extra volume of oil to thesuction side 175 of the hydraulic unit 157 to permit its operation as amotor. Sensors 195, 196 (for which there are no equivalent components inthe hydraulic control subsystem illustrated in FIG. 5) are operativelydisposed, respectively, between the solenoid operated valve 193 and thesuction side 174 of the hydraulic unit 155 and between the solenoidoperated valve 194 and the suction side 175 of the hydraulic unit 157.The pump 188/motor 189 and the solenoid operated valves 193, 194 needonly be operated when the speed of the pulley assemblies 140, 141 fallbelow a rotational rate at which the centrifugal forces exerted by thebelt engaging elements 164, 165 (FIG. 7) fall below a value at whichthey are the dominant forces.

The status of the flow rate control valves 182, 184, 194, 194 isdetermined by outputs from a control module 186 which receives inputinformation from the pressure sensors 181, 183, 195, 196 and fromexternal sources such as engine speed and throttle position. Atoperating speeds at which centrifugal force becomes a substantialfactor, when the control module 186 senses that the pressures existingacross the hydraulic units 155, 157 (and hence the transmission ratiobetween the pulley assemblies 140, 141), are incorrect for sensed engineand throttle position conditions (as, for example, during rapidacceleration), the control module responds by decreasing the flow areathrough the flow rate control valve 182 and increasing the flow areathrough the flow rate control valve 184 which momentarily slows thehydraulic unit 155 and speeds up the hydraulic unit 157 until theeffective radius of the driving pulley assembly 140 is decreased andthat of the driven pulley assembly 141 is increased to reach a new andcorrect system balance point. Also, when decelerating and using theengine as a partial vehicle brake, the control module 186 causes theflow rate through the flow rate control valve 184 to increase and thatthrough the flow rate control valve 182 to decrease, momentarilyspeeding the hydraulic unit 157 and slowing the hydraulic unit 155 untila system balance is again achieved.

As previously mentioned, when the rotational speeds of the pulleyassemblies 140, 141 fall below a value at which the centrifugal forcesexerted by the belt engaging elements 164, 165 are substantial, thehydraulic control system illustrated in FIG. 8 must revert to operationsimilar to that of the hydraulic control system illustrated in FIG. 6.This is achieved by energizing the motor 189/pump 188 and additionallygoverning the flow areas through the flow rate control valves 193, 194,to provide pressures appearing on the suction sides 174, 175 of thehydraulic units 155, 157 (as sensed by the sensors 195, 196) such thatthe hydraulic units operate as motors rather than pumps in this lowerspeed range. The result is that, in this lower speed range, thehydraulic torque input (albeit reversed from the drags discussed inconjunction with the embodiment illustrated in FIGS. 5 and 6, exerted bythe hydraulic units 155, 157) become the sole forces determining thetransient adjustment of the relative positions of the inner and outerguideway discs in the pulley assemblies 140, 141 and hence the transferspeed ratio between them.

This operation may be summarized as follows: When the rotational speedof a pulley assembly is above a value at which the centrifugal forceexerted by the belt engaging elements it carries becomes a dominantforce, the control subsystem operates the corresponding hydraulic unitto effect a change in the angular relationship between the first andsecond guideway disks to cause a tendency for change in the radialpositions of the belt engaging elements opposite to the tendency forchange in the radial positions due to centrifugal force. Conversely,when the rotational speed of a pulley assembly is below a value at whichthe centrifugal force exerted by the belt engaging elements it carriesbecomes a dominant force, the control subsystem operates thecorresponding hydraulic unit to effect a change in the angularrelationship between the first and second guideway disks to cause atendency for change in the radial positions of the belt engagingelements which are in the same direction as the tendency for change inthe radial positions due to centrifugal force.

While the principles of the invention have now been made clear inillustrative embodiments, there will be immediately obvious to thoseskilled in the art many modifications of structure, arrangements,proportions, the elements, materials, and components, used in thepractice of the invention which are particularly adapted for a specificenvironment and operating requirements without departing from thoseprinciples.

I claim:
 1. A continuously variable transmission including driving anddriven pulley assemblies coupled by a flat drive belt, each said pulleyassembly comprising:(A) a shaft; (B) a pair of pulley sheaves; (C) aseries of belt engaging elements, each said belt engaging elementhaving:1. an elongated central shank including a drive surface adaptedto be engaged by said drive belt;
 2. a first bearing region at a firstend of said central shank; and
 3. a second bearing region at a secondend of said central shank; (D) each said pulley sheave including:1. apair of relatively movable guideway disks lying alongside each other injuxtaposition;a. an inner guideway disk of each said pair including afirst series of guideways extending in one direction; b. an outerguideway disk of each said pair including a second series of guidewaysextending in a second direction; c. said first and second series ofguideways providing intersections for capturing and locating saidbearing regions of said belt engaging elements, said intersectionsproviding locations for said bearing regions to establish radialpositions of said belt engaging elements with respect to said shaft; 2.means connecting said inner guideway disks of said pulley sheavestogether to establish an inner guideway disk structure which rotatesabout said shaft;
 3. means connecting said outer guideway disks of saidpulley sheaves together to establish an outer guideway disk structurewhich rotates about said shaft;
 4. means coupling at least one of saidguideway disks to said shaft for rotation therewith;
 5. a hydraulic unitcapable of operating as a hydraulic motor and as a hydraulic pump, saidhydraulic unit having:a. an input for receiving hydraulic fluid; and b.an output for delivering hydraulic fluid;
 6. gear reduction meanscomprising a harmonic drive including:a. a wave generator coupled tosaid hydraulic unit; b. a flexspline; c. a circular spline connected toone of:i. said inner guideway disk structure; and ii. said outerguideway disk structure; and d. a dynamic spline connected to theremaining one of:i. said inner guideway disk structure; and ii. saidouter guideway disk structure; said gear reduction means differentiallycoupling:a. said hydraulic unit; b. said inner guideway disk structure;and c. said outer guideway disk structure;
 7. a control subsystem forgoverning the operation of each said hydraulic unit by:a. selectivelycontrolling the supply of hydraulic fluid to said input of saidhydraulic unit; and b. selectively controlling the discharge ofhydraulic fluid from said output of said hydraulic unit; whereby: 8.when the rotational speed of said pulley assembly is above a value atwhich the centrifugal force exerted by said belt engaging elementsbecomes a dominant force, operating said hydraulic unit to effect achange in the angular relationship between said first and secondguideway disks communicated through said gear reduction means whichcauses a tendency for change in the radial positions of said beltengaging elements opposite to the tendency for change in the radialpositions of said belt engaging elements due to centrifugal force; and9.when the rotational speed of said pulley assembly is below a value atwhich the centrifugal force exerted by said belt engaging elementsbecomes a dominant force, operating said hydraulic unit to effect achange in the angular relationship between said first and secondguideway disks communicated through said gear reduction means whichcauses a tendency for change in the radial positions of said beltengaging elements in the same direction as the tendency for change inthe radial positions of said belt engaging elements due to centrifugalforce.
 2. The continuously variable transmission of claim 1 in which, insaid driving pulley assembly:(A) said circular spline is connected tosaid inner guideway disk structure; and (B) said dynamic spline isconnected to said outer guideway disk structure;and in which, in saiddriven pulley assembly: (C) said circular spline is connected to saidouter guideway disk structure; and (D) said dynamic spline is connectedto said inner guideway disk structure.
 3. The continuously variabletransmission of claim 2 which further includes, in said controlsubsystem:(A) a hydraulic fluid pump having an output coupled to saidinput of each said hydraulic unit; (B) a solenoid-operated valve in linebetween said output of said hydraulic fluid pump and said input of eachsaid hydraulic unit; and (C) a flow rate control valve in line with theoutput from said hydraulic unit;and in which: (D) the operation of saidhydraulic unit as a motor is obtained by opening said solenoid-operatedvalve; and (E) the load of said hydraulic unit is varied by changing theflow area through said flow rate control valve.
 4. The continuouslyvariable transmission of claim 3 which further includes:(A) in saiddriving pulley assembly, first spring bias means connected to said innerand outer guideway disk structures to urge said inner and outer guidewaydisk structures toward relative movement which would move said beltengaging elements radially inwardly; and (B) in said driven pulleyassembly, second spring bias means connected to said inner and outerguideway disk structures to urge said inner and outer guideway diskstructures toward relative movement which would move said belt engagingelements radially outwardly.